An Improved Friction ModelFor Spark Ignition Engines

by

Daniel Sandoval

Submitted to the Department of Mechanical Engineeringin Partial Fulfillment of the Requirements for the Degree of

Bachelor of Science in Mechanical Engineering

at the

Massachusetts Institute of Technology

May 2002

@ 2002 Daniel SandovalAll rights reserved

MASSACHUSETS INSTITUTEOF T ECHNOLOGY

JUN 17 2003

LIBRARIES

The author hereby grants to MIT permission to reproduce and to distributepublicly paper and electronic copies of this thesis document in whole or in part.

Signature of Author.................................Department of Mechanical Engineerihg

May 10, 2002

I

Certified by....................................................................... ............ j.........John B. Heywood

Sun Jae Professor of Mechanical EngineeringThesis Supervisor

A ccepted by ...................... ................... ..................................................Ernest Cravalho

Chairman of the Undergraduate Thesis CommitteeDepartment of Mechanical Engineering

ARCHIVES

AN IMPROVED FRICTION MODEL

FOR SPARK IGNITION ENGINES

by

Daniel Sandoval

Submitted to the Department of Mechanical Engineeringon May 10, 2002 in partial fulfillment of the

requirements for the Degree of Bachelor of Science inMechanical Engineering

Abstract

The details of a modified model that predicts friction mean effective pressure (fmep) for spark-ignitionengines are described. The model, which was based on a combination of fundamental scaling laws andempirical results, includes predictions of rubbing losses from the crankshaft, reciprocating, and valvetraincomponents, auxiliary losses from engine accessories, and pumping losses from the intake and exhaustsystems. These predictions were based on engine friction data collected between 1980 and 1989. Someof the terms are derived based on lubrication theory. Other terms are derived empirically frommeasurements of individual friction components, rather than the basic processes themselves. Currentengine developments (improved oils, surface finish on piston liners, valve train mechanism) suggestedthat the model needed modification to predict modem engine losses due to friction. Modifications in oilviscosity, piston ring tension and gas pressure contribution on piston assembly, liner roughness, andvalvetrain mechanism were made. The sum of the predictions now gives reliable estimates of spark-ignition engine fmep. The inclusion of oil viscosity scaling with temperature results in cold enginefriction predictions about twice the value for warmed up engines. This agrees with the limited enginefriction data used to modify the friction model.

Thesis Supervisor: John B. HeywoodTitle: Sun Jae Professor of Mechanical Engineering

Director, Sloan Automotive Laboratory

2

11100 ! _W , ,

Contents

1 Introduction 7

2 Engine Friction 8

2.1 Components of Engine Friction ............................................. 8

2.2 Measuring Friction .................................................... 8

3 Prior and Improved Friction Model 9

3.1 Background.......................................................... 93.2 M odification of Friction M odel .............................................. 113.3 Component Friction Models ............................................... 12

3.3.1 Component Breakdown ............................................... 123.3.2 Crankshaft Friction ................................................. 133.3.3 Reciprocating Friction .............................................. 133.3.4 Valvetrain Friction................................................ 153.3.5 Auxiliary Friction.............................................. 163.3.6 Pumping Losses ................................................. 17

3.4 Comparison of Prior and Improved Model................................... 18

4 Results 19

4.1 Engine Parameters ......... ........................................... 194.2 Crankshaft Friction .... ............................................... 194.3 Reciprocating Friction...............................................21

4.3.1 Terms without Gas Pressure Loading....................................214.3.2 Terms with Gas Pressure Loading..................................... 21

4.4 Valvetrain Friction........................................ .......... 214.5 Auxiliary Friction.................................................. 224.6 Total Mechanical Losses................................................264.7 Pum ping Losses ............ .............. ................................ 294.8 Engine Warm Up.................................................... 30

5 Summary 32

Bibliography 33

Appendix 34

3

List of Figures

3-1 Stribeck Diagram - Lubrication Regime. Figure from [4] ................................ 10

3-2 Stribeck Diagram - Lubrication Regime. Figure from [5] ............................... 10

3-3 Friction Data for 5.4L Engine at 2500 rpm - WOT .................................... 18

4-1 Comparison of crankshaft friction for 3.OL engine ..................................... 20

4-2 Comparison of crankshaft friction for 5.4L engine ..................................... 20

4-3 Comparison of reciprocating friction w/out gas for 3.OL engine ........................... 23

4-4 Comparison of reciprocating friction w/out gas for 5.4L engine ........................... 23

4-5 Comparison of valvetrain friction for 3.OL engine ..................................... 24

4-6 Comparison of valvetrain friction for 5.4L engine ..................................... 24

4-7 Comparison of accessory friction for 3.OL engine ..................................... 25

4-8 Comparison of accessory friction for 5.4L engine ..................................... 25

4-9 Comparison of complete mechanical friction for 3.OL engine ............................. 26

4-10 Comparison of complete mechanical friction for 5.4L engine ............................. 27

4-11 Comparison of complete mechanical friction for 6.8L at 55 kPA MAP..................... 27

4-12 Comparison of total mechanical friction with piston gas loading for 3.OL engine ............. 28

4-13 Comparison of total mechanical and gas exchange losses for 5.4L engine ................... 28

4-14 Comparison of pumping losses for 6.8L engine ...................................... 29

4-15 Friction for varying temperature at 2000 rpm for 3.OL engine............................. 31

4-16 Friction for varying temperature at 2000 rpm for 5.4L engine............................. 31

A-i Cross-section of an engine at intake stroke .......................................... 35

A-2 Modified Expression for Total FMEP .............................................. 36

4

imillbMMM a ele Illllo Molimillllin III M7". deliumma dulillHinaan a

List of Tables

3.1 Typical Test Data for Oil Viscosity.................................................. 11

4.1 Engine Param eters............................................................... 19

4.2 Constants for Valvetrain Mechanism Terms...........................................22

A .1 D efinition of Sym bols ............................................................ 34

A .2 O il G rades......................................................................35

A.3 Sample Input Configuration for Friction Model........................................37

5

Acknowledgements

I would like to thank Professor Heywood for giving me an opportunity to do research in engines.

It was a great opportunity to work with such a knowledgeable patient, and supportive advisor. There

were numerous times that he took the time to stay that extra time necessary to clarify any questions I had

and to make sure I kept on going.

I would like to thank my parents. There are no words that can describe my appreciation for their

love and the sacrifices they have made so that I can pursue my goals and aspirations. I have been

extremely fortunate to have them in my life and thank them for their endless support and encouragement.

I thank everyone in the Sloan Automotive Lab, always willing to help me find information or

understand certain concepts. In particular, Jenny Topinka, graduate student, and Dr. Tian Tian, for

giving me guidance and advise throughout my work.

Thanks to the professional engineers in the automotive industry that guided me in understanding

and interpreting the engine data.

6

1 Introduction

There is strong interest in improving spark-ignition and in reducing fuel consumption. For these reasons,

it is of great importance to have accurate predictions of mechanical losses, a great part of which are

friction losses due to the relative movement of adjacent components of the engine. Numerous

publications have been made that detail measured and analysis based frictional losses incurred when

running reciprocating engines. However, results are difficult to validate since the total friction loss is the

summation of losses arising from many components in the engine. These friction components respond

differently to variations of pressure, temperature, and speed as engine load varies [1].

The model for predicting "friction mean effective pressure" (fmep) described in this paper was

developed to predict spark-ignition engine friction using basic engine design and operating parameters as

inputs [2]. However, the data used to develop this model date back to the late 1980s. In order for this

model to predict friction loses that occur in current engines, it was necessary to compare it to current

engine friction data. This allowed an assessment and a recalibration of the friction terms in the model.

Current engine friction data was used to determine the changes that needed to be made to certain

constants in the model. The model was also expanded to include the lubricant viscosity in the model to

include the effects of component temperature on the engine during cold start and warm-up transients.

Finally, a comparison between pumping loses in the intake, exhaust, and the terms representing changes

in cylinder gas pressure were made. The effect of these latter changes was minimal.

This thesis is organized as follows. A brief overview of the components of the engine friction

model is presented, to explain what considerations were included in developing a friction model. The

prior and improved models are then presented to show and explain the modifications that have been made

to the friction terms. Data are then presented to compare the "old" and "new" model in a brief summary

and the in a comparison of model breakdown against current engine friction data. Developing this model

further, also allowed the effects of changing coolant temperature to examine the model effect of changing

from start-up to warm-up conditions. Finally, a summary is presented to explain the improvements in the

model and suggest some changes to keep it robust.

7

2. Engine Friction

2.1 Components of Engine Friction

Engine friction losses can be divided into three main categories: mechanical or rubbing losses, pumping

losses, and auxiliary component losses. These losses are defined as follows [3]:

1. Rubbing losses or mechanical losses are those losses which result from relative motion between

solid surfaces in the engine; i.e., the motion between a piston and a cylinder wall or a crankshaft

journal and a main bearing. Relative motion does not require that the two solids be in contact

with each other; in fact, it is generally the case that there is a film of lubricant between the

surfaces.

2. Pumping losses are those losses associated with transporting fluids through the cylinder - made

up of intake and exhaust pumping.

a. Intake - Drawing the gas mixture through the intake system and into the cylinder.

b. Exhaust - Expelling the burned gases from the cylinder and out of the exhaust system.

3. Auxiliary component losses are losses in driving engine accessories. These can include: the fan,

the water pump, the oil pump, the fuel pump, the generator, a secondary air pump for emission

control, a power-steering pump, and an air conditioner. Under most engine tests, the main

accessories considered are those necessary for normal engine operation: the oil pump, the water

pump, the fuel pump, and sometimes the alternator.

2.2 Measuring Friction

Two commonly used methods to obtain engine friction are "firing" and "motoring" of an engine.

A true measurement of friction in a firing engine can be obtained by subtracting the brake mean effective

pressure (output) from the indicated mep (energy transfer to each piston) determined from accurate

measurements of cylinder pressure throughout the cycle. Pumping mep is obtained from the p-V data

leaving mechanical and auxiliary friction losses. The second method is direct motoring of the engine, the

engine is driven without firing, under conditions as close as possible to firing; for example, similar

coolant temperature [3]. By motoring an engine and sequentially dismantling it, each component of

mechanical friction contribution can be determined.

The motoring losses are different from the firing losses in terms of the gas loading on the piston,

the piston and cylinder temperatures, and the exhaust blow down phase in the exhaust stroke. The lower

gas loadings during motoring lower the rubbing friction. The temperatures are lower in a motoring test.

Also, the gases in the exhaust phase are higher in density than under firing conditions [3].

8

3 Prior and Improved Friction Model

3.1 Background

The total engine friction prediction was defined as the sum of the predictions for mechanical, auxiliary,

and pumping losses [2]. The rubbing and auxiliary losses were expressed as friction mean effective

pressures, and the pumping loss was expressed as a pumping mean effective pressure, where mean

effective pressure, mep, is defined as the work per cycle per unit of displaced volume:

Wmep- (1)

Vd

For individual rubbing friction interfaces and the auxiliary components, terms that related fmep to the

basic engine design and operating parameters were developed. For the pumping loss, terms that related

intake and exhaust system pressure drops to the appropriate design and operating parameters were

developed. The resulting model was in the form:

tfmep = mfmep + afmep + pfmep (2)

Relating rubbing interface fmep to design and operating parameters was a three-step process.

First, an assumption that identified the type of lubrication present was made to determine the relationship

between the friction coefficient and a dimensionless duty parameter, which was a function of viscosity

(g), velocity (V), and unit load (P). Figures 1 and 2 show the Stribeck diagram that plots friction

coefficient vs. duty parameter, gV/P. There are three lubrication regimes that are important for engine

components - boundary, mixed, and hydrodynamic lubrication. In boundary lubrication (no apparent

lubrication) the friction coefficient is essentially constant. In mixed lubrication (some lubrication) the

friction coefficient was assumed to vary inversely with engine speed. In hydrodynamic lubrication (full

film lubrication) the friction coefficient was assumed to vary with a term proportional to the duty

parameter [2].

The second and third steps in the rubbing friction term development were to use the friction

coefficient to derive a term proportional to fmep for the interface in consideration and to "calibrate" the

terms by multiplying them by constants based on empirical results. Patton explained the detailed analysis

used to derive the terms, and used friction data for individual engine components and groups to calibrate

these terms.

The auxiliary and pumping fmep terms were independent of lubrication. These terms are

explained in Section 3.3.1, Component Breakdown.

9

LubricantViscosity

- W. Mean SurfaceL-4 Roughness

0FI

FE ----- Effective

Gil FilIM

1:-2,

RI R2

DUTY PARAMETER pV/P

Variable Description:FB Coefficient of friction for boundary lubrication.RI Ratio of effective oil film to surface roughness when boundary lubrication begins.R2 Ratio of effective oil film to surface roughness when mixed lubrication begins.

0.1

Coefficlent 0 0 1of friction

0.001

0.001 0.01 0.1 1.0 10

HydrodynamicBoundary Mixed-

-

* I uA

Thin film i

Piston rings

r mn, si" r -

Engine bearings

Thick f

p y viscosityv sliding velocityp bearing pressure

Sommerfield Number x 104

Figures 3-1 and 3-2: Stribeck Diagram -Lubrication Regimes. Figures from [41 and [51

10

0.00,

ilm

ZI

3.2 Modification of Friction Model

The purpose of this investigation was to determine how well the friction model predicted friction for

current engines to then improve the predictive accuracy of the model. Engine development has

significantly improved engine performance and efficiency over the past 10 years or so. Modifications in

oil viscosity, crankshaft bearings and seals, piston design, piston ring design, liner roughness, and

valvetrain mechanism, have decreased the total amount of friction losses in modem engines [1].

The improvements made in the friction model were made knowing that the friction work

components fall under three main categories: those that are independent of speed (boundary friction),

proportional to speed (hydrodynamic friction), or those proportional to speed squared (turbulent

dissipation). Other components are a combination of these [3].

Literature on engine lubrication suggested that the hydrodynamic terms in the model should be

modified to compensate for the differences between oil grades and the temperature dependence of engine

oil viscosity. The fundamental literature for tribology of reciprocating engines shows that the viscosity

scaling that should be included in the hydrodynamic friction terms has the form [6]:

p (T )Yscaling = (3)

p , (T, )

where g(T) is the viscosity of the oil in the engine for which friction predictions are being made, and

po(T.) is the reference viscosity for the engines that provided the data used to calibrate the model when it

was first developed. Patton's paper [2] indicates that a reference 10W30 oil was used (viscosity 10.6 cSt

at 90'C) in obtaining the engine data he used. This falls in the midrange of current oil grades used in

engines. Table 1, shows current typical viscosity data for different oil grades:

Table 3.1: Typical Test Data for Oil Viscosity [7]

Oil Grade 5W-20 5W-30 1OW-30 1OW-40 20W-50CPS Number 220135 220013 220019 220059 220060Viscosity, KinematiccSt at 400C 49.2 66.1 74.8 98.9 174.4cSt at 1000C 8.6 11 10.8 14.4 19.1

A more detailed set of oil grades that listed kinematic viscosity v is given in the Appendix A.2. Since the

relationship between v and p is a scaling factor, which is constant density, Equation 3 was written as:

'scaling = Fv(T) (4)v0 (T0)

11

m Illiallellilik liilllalM Enonnoll ibl3 id-rm. alilliisil i r / .

where v(T) is the viscosity of the oil in the engine that is being tested, and v0(T) is the reference viscosity

at reference temperature T.. The Vogel equation, which gives the relation between temperature and low

shear kinematic viscosity, was used in the form [8]:

vo = k exp{1 (5)

where vo is the kinematic viscosity of the low shear rate oil in cSt, and k(cSt), 6 1(*C) and 02('C) are

correlation constants for an oil and T is the oil temperature in 'C. Furthermore, the low shear rate

viscosity was multiplied by a ratio of pt/g. to convert to the high shear rate kinematic viscosity used in

the model. This was done because most of the engine components operate at a high shear range where

multi grade oils exhibit shear thinning.

V= (6)

The oil grades and constants used are listed in the Appendix A.2 [8].

Modifications in the boundary and mixed lubrication terms were also made. The terms in each

component of the friction model that were changed, and the explanation for that change, are detailed in

the following section.

3.3 Component Friction Models

The following sections describe each term in the crankshaft, reciprocating, valvetrain, auxiliary, and

pumping models. Each section describes the lubrication regime that was used to develop the terms for

each model and the modifications made to each term. A detailed explanation on the derivation of each

term can be found in Patton [2]. The definitions of all symbols and the expression for the total fmep are

included in the Appendix A. 1 and A-2.

3.3.1 Component Breakdown

1. Rubbing friction was divided into three component groups:

Crankshaft - Main bearings, front and rear main bearing oil seals.

Reciprocating - Connecting rod bearings, pistons, and piston rings.

Valvetrain - Camshafts, cam followers, and valve actuation mechanisms.

2. Pumping losses were:

Intake system, intake and exhaust valve(s), and exhaust system.

Turbulent dissipation in hydrodynamic journal bearings (in rubbing friction component).

12

3. Auxiliary losses were:

Oil pump, water pump and alternator (necessary for normal engine operation).

* A cross sectional view of an engine showing the components is found in the Appendix A-1.

3.3.2 Crankshaft Friction

Crankshaft friction was estimated by adding a prediction of front and rear main bearing seal friction to a

prediction of main bearing friction. The main bearing friction prediction included a hydrodynamic

journal bearing term and a turbulent dissipation term, the latter accounting for losses due to the transport

of oil through the bearings. The three terms that make up the crankshaft friction were:

D 104ND'L ng D 2N 2 ncfinep = 1.22 X105 +3.03 x10~4 +1.35 x10-'o (7.a)

(B 2 Snc B 2Snc nc

The first term gave the friction of the main bearing seals. The seals were assumed to operate in the

boundary lubrication regime due to direct contact between the seal lips and the crankshaft. The seal lip

load was assumed constant. The second term was for the main bearing hydrodynamic friction - derived

using the friction coefficient for hydrodynamic lubrication. The last term accounted for the turbulent

dissipation, the work required to pump fluids through flow restrictions [3].

Modifications

The only modification made to the crankshaft friction model was to include the viscosity scaling in the

hydrodynamic friction term.

Dp ND Lbng D 2N 2 ncfinep = 1.2 2 x105 j + 3.03 x10- 4 b b + 1.35 x10-1o bNb (7.b)(B 2Snc pITO B 2SnC nc

3.3.3 Reciprocating Friction

The reciprocating component group friction prediction included piston, piston ring, and connecting rod

friction. Piston ring friction was divided into two terms; one that predicted friction for the piston rings

without gas pressure loading, and one that predicted the increase in piston ring friction caused by gas

pressure loading. The resulting friction terms were:

Terms without gas pressure loading

13

rfmep = 2.94 x102K jJ+ 4.06 x104 1 + -I + 3.03 x10-4 ND4nb (8.a)BN B 2 B 2Sne

The first term gives the piston friction assuming hydrodynamic lubrication. The second term is for the

piston ring friction term developed assuming mixed lubrication. The function, 1+1000/N, was selected to

make the friction coefficient decrease by a factor of 2.5 from low to high speeds. The last term accounts

for the hydrodynamic journal bearing fmep term for connecting rod bearings. This term is the same as the

term for the main bearings in Equation 7.a.

Terms with gas pressure loading

rfinepgas = 6.89- [0.088r, +0.182r (1.33-KSP) (9.a)Pa

The term for piston ring friction due to gas pressure loading used the product of intake pressure (p) and a

factor which included compression ratio (re) raised to constant - relating to the physics of the

compression process.

Modifications

There were several modifications made to the piston terms, as there have been considerable

improvements in piston rings and liner to reduce the amount of friction between these two surfaces.

Terms without gas pressure loading

rfmep = 2.94+102 #A('J+4 . 6 x14 ' C, + 2+ 3 .03xi1-4 - ND Ln (8p B (F, ) N B2 ) To ( B2 Sn (

For the equation without gas pressure effect, the viscosity scaling was incorporated into the piston friction

and connecting rod bearing hydrodynamic terms. For the piston ring friction term, three modifications

were made. Two factors, piston ring tension and surface roughness, were included to take into account a

decrease in piston friction due to these two terms. Although studies have been made to measure these

contributions, the improvements can only be accounted on a case-by-case basis. For this study, the

engine data for piston friction was used to give a value to these terms. The last change was to decrease

the value of the function, 1 +1000/N to I +500/N, to make the friction coefficient decrease by a factor of

1.8 instead of 2.5, which affects the friction coefficient for boundary lubrication.

14

Terms with gas pressure loading

rfinepgas = 6.89 0. 088 r + 0.182 C F' r ( 2KSP) (9.b)

For the equation with gas pressure effect, the viscosity scaling was included in the first term. For the

second term, the piston ring tension factor was also included. For the exponential term, the constant K

was doubled to reflect a reduction in liner roughness.

3.3.4 Valve Train Friction

The valvetrain component friction prediction included estimates of camshaft, cam follower, and valve

actuation mechanism friction for a variety of valvetrain configurations. Flexibility dictated that the model

be general enough to predict friction for multiple types of interfaces, yet selective enough to include only

the effects of specific components for a desired valve configuration. The terms modeled camshaft bearing

hydrodynamic friction, cam follower friction for flat or roller followers, oscillating hydrodynamic

friction, and oscillating mixed lubrication. The resulting terms were:

Nnb 1000 nV Nnv L C("N 05n, 1000) Inv (Oavfmep=244 +C 1+ IJ +C, +Co h +C0 ,(1+ ' (.B2Sn, f N )Sn. n BSn, N Sn.

The first term gave the camshaft bearing hydrodynamic friction. This term had a form similar to those of

the main and connecting rod journal bearing terms. Engine data indicated that some of the friction was

independent of piston speed (extrapolation back to zero engine speed gave a nonzero fmep value), so a

constant 4.12 kPa was added to the camshaft bearing term. The constant value represented the boundary-

lubricated friction due to the camshaft bearing seals. The next two terms predicted friction resulting from

relative motion between the cam lobe and the cam follower. The second term predicted friction in the

mixed lubrication regime for flat follower configurations. The third term predicted rolling contact friction

for roller follower configurations. The fourth term, oscillating hydrodynamic friction, predicts friction

caused by relative motion between relative motion between valvetrain components whose lubrication

states were either completely or partially hydrodynamic, such as the valve lifter in the lifter bore or the

valve in the valve guide The constants for the valvetrain terms (Cf Cf Cho, Corn ) were dependent on the

type of valvetrain configuration being modeled. The values of the constants for SOHC-finger flat and

15

roller follower configurations were obtained using engine data. The values for the constants for other

valve train configurations were estimated as fractions of the SOHC constants.

Modifications

There were two main modification made to the valve train friction terms. The first was including the

viscosity scaling in the camshaft bearing and oscillating hydrodynamic friction terms. The second, was to

lower the value of the function, 1 +1000/N to 1 +500/N, in the flat follower and oscillating mixed

lubrication term.

vfnep=244 - +C 1+ +, + +N + C J + CO, 1+ (10.b)p,0 B2n N S f Sn, - BSn, N Sn,

3.3.5 Auxiliary Friction

The auxiliary component friction prediction was an estimate of the sum of oil pump, water pump, and

non-charging alternator friction. All of the auxiliary component fmep were assumed to be proportional to

engine displacement. The auxiliary fmep model was:

aftnep = 6.23+5.22x10- 3 N -1.79 x10-7 N2 (1

The constants were determined from a group of small, high-speed diesel engines. The negative

coefficient for the engine speed squared term resulted from a gradual leveling off of oil pump fmep;

typically, as oil pressure reaches its limited value in the middle speed range, less work is required of the

pump.

Modifications

There were no modifications made in the auxiliary components losses since the terms were only

dependent on engine speed. Also, there was not enough information available to determine properties of

the components being tested in the first set of engines or in the engine data used to modify the model.

16

- Eft Am.sri iMWW -i.. 11lis nme llh v - | Iusn - -

3.3.6 Pumping Losses

The pumping loss model predicted intake and exhaust pumping mean effective pressure (pmep).

Pumping work for either the intake or exhaust strokes was defined as the difference between cylinder

pressure and atmospheric pressure, integrated over the volume of the intake or exhaust stroke. The intake

and exhaust terms are listed below.

Intake

pmep i = (p P + 4.12 x 10 -3 (12)p n, r

The first term for the intake pmep was assumed to be equal to the sum of the intake manifold vacuum and

the intake valve pressure drop term. Intake manifold vacuum was calculated directly as the difference

between atmospheric and intake pressure (pa - pi). The second term measured intake valve pressure drop.

Exhaust

(P 2)2

/2

pmepe = 0.178 - S, + 4.12 x 10 -3 P 2 P , (13)(P a Pa n 2r,

The first term in the exhaust model measured the exhaust system pressure drop. It was derived to

measure typical production engine systems. The second term measured exhaust valve pressure drop.

Modifications

The constant for both valve pressure drop terms in Patton's model was based on results for diesel engines.

The assumption that had been made was that pmep for SI engines was similar to pmep for diesel engines.

Results from motoring tests of SI engines have indicated that the constant may have been too high. In

this modified model, the constant for the valve pressure drop was changed to 3.0 x10 3 kPa-s 2/m 2,

following Patton's suggestion. The smaller constant may also take into account improvements in intake

manifold design. With this change, the model gave a more accurate prediction for the intake and exhaust

losses.

17

1BEEK1E"''"~ 11 m 1 11II ilMoisilromminilllonll INTT" . - - ||Billl -

3.4 Comparison of Prior and Improved Model

Figure 3-3, shows the friction values for each friction component model for the 5.4L engine. The last

three columns show the total mechanical, pumping, and total friction mean effective pressure. The total

mechanical column was the sum of the first five columns and the total friction column included the

pumping losses. The modified total friction model expression was included in Appendix A-2.

M Prior Model S Improved Model

160

140 -

120 -

100 -

80

60 -

40 -

20 -

0-crankshaft piston piston/gas valvetrain auxiliary tmmep pmep

Figure 3-3: Friction Data for 5.4L Engine at 2500 rpm - WOT

No large differences can be seen in the mechanical friction models (crankshaft, piston, piston/gas, and

valvetrain) since the viscosity scaling was only included in the hydrodynamic terms - which make a small

contribution to the total value of friction for each of these models. However, there is a difference of about

6 kPa in the piston/gas term due to the piston ring tension, surfaces roughness, and mixed lubrication

factors that were included in that model. The auxiliary losses were the same since the model was not

modified. For the pumping model, the change made to the valve pressure drop constant decreased the

pumping loss distribution by 25 kPa.

18

tfmep

'

I

4 ResultsSection 4.1 lists the three engines (5.4L, 3.OL, and 6.8L) that were used to validate and compare the

changes that were made to the friction model. The first two engines were used to test the mechanical

friction model. The last engine was used to test the pumping losses model. A comparison between the

model and data for the 5.4L and 3.0L engine is presented to explain each friction model. The prototype

engine was used to compare the pumping loss model at varying intake pressures. The rest of the sections

will show the results for the improved friction model.

4.1 Engine Parameters

The modifications to the friction model were based on friction data from two current production engines

and a prototype engine. The two current production engines were tested by motoring. This gave

component friction data that allowed us to modify and compare the results for each component of the

model. The engines were tested at an engine coolant temperature of 90 0C. The following table lists the

engine parameters that were used in the friction model.

Table 4.1: Engines Parameters

Engine Type Current Current Prototype

Engine 5.4L 3.OL 6.8L

Engine Information

Layout/ # Cylinders V8 V6 V10

Valves/Cylinder 2 4

Valvetrain Type SOHC DOHC SOHC

Valvetrain Mechanism Direct Acting Roller Finge Roller Finger

Bore/Stroke Ratio 0.85 1.1 0.9

Compression Ratio 9.0 10.0 11.75

4.2 Crankshaft Friction

Figures 4-1 and 4-2, show that the improved friction model gave the same trend but significantly under

predicted the crankshaft friction measurement for both engines. The crankshaft terms may oversimplify

the complexity of modeling the crankshaft friction, since about 80% of the friction was accounted for by

the hydrodynamic and turbulent dissipation terms. It also may be that testing the crankshaft without the

pistons and connecting rods significantly increased the friction since the counter weights offset the

balance of the crankshaft under these conditions.

19

-U-- Data -Crankshaft -*-Model -Crankshaft

1000 1500 2000 2500 3000 3500 4000 4500 5000

rpm

Figure 4-1: Comparison of crankshaft friction for 30L engine

-U- Data -Crankshaft -*- Model -Crankshaft

1000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000

rpm

Figure 4-2: Comparison of crankshaft friction for 5.4L engine

20

30

25

20

15

10

5

0

30

25

20

15

10

0

0

4.3 Reciprocating Friction

4.3.1 Terms without Gas Pressure Loading

Figures 4-3 and 4-4, show the friction losses for the piston group (piston, piston rings, and connecting

rod). The model for the 3.0L engine matched well during the low speed range and then fell a consistent

amount below the data. At high speeds, the difference between the model and data was about 13 kPa.

For this model, the piston ring hydrodynamic term dominated the friction contribution. The steep

increase in the friction increased with speed, Equation 8.b, due to increasing piston velocity. The slight

decrease in the data trend may have occurred due to the engine oil characteristics that were used when

motoring the engine. The model for the 5.4L engine matched extremely well from low to high speed

ranges. Again, the piston ring hydrodynamic term dominated the friction contribution as the friction

increased due to the increase in mean piston speed.

This reciprocating model showed the similar increasing trend for the piston group friction. The

data did not show such a trend. In the model, it is difficult to predict the dynamics of the tangential forces

and lubrication characteristics at varying piston speeds. The model and data for the 5.4L engine showed

the friction trend at various piston speeds, with an upward turn at 6000 rpm for the data.

4.3.2 Terms with Gas Pressure Loading

Data was not available to compare the gas pressure loading effect on the piston friction term because data

has to be collected with the cylinder head and valvetrain mechanism. Figures 4-12 and 4-13, explained in

section 4.6, include the gas exchange losses for the 3.0L and 5.4L engine.

4.4 Valvetrain Friction

Figures 4-5 and 4-6, show valvetrain friction losses for the 3.0L and 5.4L engine, respectively. In the

3.0L engine, the model over predicted the valvetrain friction, while in the 5.4L engine, the model under

predicted the valvetrain friction. However, both models did maintain a similar trend to the engine data.

The match between model and data fit well considering the flexibility of that the model had to predict

friction for multiple types of interfaces. Table 4.2, shows the constants available for the different

valvetrain mechanisms used in engines. A set of constants was added to the original table to be able to

model the 3.0L engine. Existing constants were used to compare the 5.4L engine data.

The constants for the 3.0L DOHC finger follower valvetrain mechanism were obtained by

matching the model to the engine data. The first three constants terms were kept the same as the

21

constants for a SOHC finger follower configuration. The oscillating mixed constant was set as 0.6 times

the constant for the SOHC finger follower. This gave the best fit between data and model.

For the 5.4L engine, the difference at low and high engine speeds in the model may have been

accounted for by the decrease in the boundary lubrication function term of 1 +1000/N to 1 +500/N.

However, the data between the 2000 to 5000 rpm range matched extremely well.

Table 4.2: Constants for Valvetrain Mechanism Terms

TYPE Cam Follower Oscillating OscillatingFlat Roller Hydrodynamic Mixed

SOHCfinger follower 600 0.0227 0.2 42.8

SOHC 400 0.0151 0.5 21.4rocker arm

SOHC 200 0.0076 0.5 10.7direct acting

DOHC* 600 0.0227 0.2 25.8finger follower

DOHC 133 0.0050 0.5 10.7direct acting

OHV 400 0.5 0.5 32.1

*Determined to have improved model match engine data for DOHC finger follower valvetrain mechanisms.

4.5 Auxiliary Friction

Figures 4-7 and 4-8, show the accessory losses for the 3.OL and 5.4L engine, respectively. The auxiliary

losses model indicates that there should be no differences between engines since accessory losses were

only based on engine speed. The model for both engines fits the data well at speeds below 3500 rpm. For

higher speeds, there is a sudden increase in the auxiliary losses based on the data. Although the

characteristics of the accessories (oil pump, water pump, and alternator) were not studied, there are many

friction losses that could have increased at higher speeds. There may have been high friction of the

bearing and seals in the accessories at higher speeds. Also, the high hydraulic power that the oil and

water pump produce, leads to an increase in the friction losses. For example, high-pressure losses occur

in the narrow cross sections that the oil and water have to flow through.

22

-- Data - Piston + Model - Piston

70

60

50

40

30

20

10

01000 1500 2000 2500 3000

rpm

3500 4000 4500 5000

Figure 4-3: Comparison of reciprocating friction w/out gas for 3.OL engine

-- Data - Piston -*- Model - Piston

120

100

80

60

40

20

01000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000

rpm

Figure 4-4: Comparison of reciprocating friction w/out gas for 5.4L engine

23

-- 40

- - Data - Valvetrain -+- Model - Valvetrain

35-

30-

15-

10-

5-

0

--a- Data - Valvetrain -$- Model - Valvetrain

12-

10-

8-

0. 6-

4-

2

01000 1500 2000 2500 3000 3500

rpm

4000 4500 5000 5500 6000

Figure 4-6: Comparison of valvetrain friction for 5.4L engine

24

1000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000

rpm

Figure 4-5: Comparison of valvetrain friction for 3.OL engine

...........

............................ ........ 77-710 7

I

jIlll,

+ Data - Accesories -+- Model - Accesories

1000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000

rpm

Figure 4-7: Comparison of accessory friction for 3.OL engine

- - Data - Accesories + Model - Accesories

1000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000

rpm

Figure 4-8: Comparison of accessory friction for 5.4L engine

25

45

40

35

30 1

25 -0.x 20-

15-

10 -

5

0

45

40

35

30

25a.2 20

15

10

5

0

4.6 Total Mechanical Losses

Figures 4-9 and 4-10, show the total mechanical friction obtained for the stripped engine tests for the 3.OL

and 5.4L engines. This is the sum of the friction values for the crank train (crankshaft, piston group

without gas pressure loading, connecting rod bearings), the valvetrain and accessories (oil pump, water

pump, and unloaded alternator). The sum of the over and under prediction of the individual components

gave a close match to the data. This makes the model suitable to predict total mechanical losses.

Figure 4-11, shows the result of the total mechanical friction obtained for the 6.8L prototype

engine, tested under firing conditions. The model over predicts the mechanical friction by about 15 %,

suggesting that continuing engine developments are leading to lower friction engines.

Figure 4-12, shows the result of the total mechanical friction with piston loading for the 3.OL

engine. The model over predicted the frictional losses by an average of 15 kPa from low speeds to high

speeds. However, both data sets follow the same trend. Again, the over and under prediction of the

model for the mechanical losses may contribution to a total over prediction for the losses.

Figure 4-13, shows the result of the of the total friction losses for the 5.4L engine, including the

intake and exhaust mean effective pressure. The model under predicts the data by about 10 kPa from low

to high speeds, but follows a close trend with increasing speed. For this engine, the model gave an

accurate match to the total actual friction losses. This gives an indication that the modifications made to

the friction model reduced any over predictions that the previous model may have given.

- Data -Complete Mechanical Friction -$- Model - Complete Mechanical Friction

180.000

160.000

140.000-

120.000-

100.000-a.

80.000

60.000

40.000-

20.000

0.0001000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000

rpm

Figure 4-9: Comparison of complete mechanical friction for 3.OL engine

26

I- Data - Complete Mechanical Friction -+-Model - Complete Mechanical Friction

200.000

180.000

160.000

140.000-

120.000

. 100.000

80.000

60.000

40.000

20.000

0.0001000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000

rpm

Figure 4-10: Comparison of complete mechanical friction for 5.4L engine

--- Data - Complete Mechanical Friction ---- Model - Complete Mechanical Friction

90.000

80.000

70.000

60.000

50.000

20.000

40.000

20.000

10.000

0.0001000 1500 2000

rpm

2500 3000

Figure 4-11: Comparison of complete mechanical friction for 6.8L at 55 kPA MAP

27

-a- Data - Motored w ith Piston Gas Loading -+- Model - Motored w ith Piston Gas Loading

200

180

160

140-

120

oL 100

80 -

60-

40

20-

01000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000

rpm

Figure 4-12: Comparison of total mechanical friction with piston gas loading for 3.OL engine

U0.

-I- Data - Motored w ith Gas Exchange Losses --- Model - Motored w ith Gas Exchange Losses

240

200

160-

120-

80

40

01000 1500 2000 2500 3000 3500 4000

rpm

Figure 4-13: Comparison of total mechanical and gas exchange losses for 5.4L engine

28

4.7 Pumping Losses

Figure 4-14, shows the pumping losses for the 6.4L engine. This engine was used to verify the intake and

exhaust friction mean effective pressure. The total pumping losses are the sum of the intake and exhaust

system, and intake and exhaust valve pressure drop suggests the model oversimplifies the complexity of

the intake and exhaust system. For this reason, the data was plotted for increasing intake map, mean

atmospheric pressure, at constant speeds. The model shows the same trend for the pumping losses for

increasing speeds and increasing map. At 1000 rpm, there is the largest difference between the model and

the actual data. For increasing speeds, this difference decreases significantly and for 3000 rpm, the model

matches the data fairly well. Also, at speeds higher than 2000 rpm, there is a slight kick in the model

prediction caused by the pressure and mean piston squared terms. At higher speeds and maps, the exhaust

system and both valve pressure drop terms have a higher contribution to the pumping losses.

10 20 30 40 50 60 70 80 90

MAP (kPa)

Figure 4-14: Comparison of Pumping Losses for 6.8L engine

29

80.00

70.00

60.00

50.00wU.

40.00

30.00

20.00

0

4.8 Engine Warm Up

The viscosity scaling included in this model allowed us to see the effect of oil temperature on the different

friction models. Shayler [10] included a modification in the previous Patton model to account for the

increased friction losses, which arise during engine warm-up. These losses were attributed primarily to

the effect of temperature on oil viscosity. In the test, the engine was allowed to soak down to room

temperature (20'C) prior to each fired friction test. A plot of friction power loss against time from engine

start was obtained. Several similar tests were carried out and all of these indicate that the friction power

loss at engine start-up is over two times higher than the hot engine friction power loss.

With the modified friction model, data could now be collected given a particular temperature to

calculate oil viscosity. Figures 4-15 and 4-16 show data that the model gave for different temperatures

which affected the oil viscosity and therefore affected the friction terms. Comparing the total mechanical

friction at 20 and 90*C shows a friction loss of about 2.1 times the hot engine friction power loss. This is

comparable to the data that Shayler obtained from the firing tests.

30

Total Mechanical -A- Recip no/gas -*- Reci p gas -U- Crankshaft

200.0

180.0

160.0

140.0

120.0-

100.0

80.0

60.0-

40.0-

20.0-

0.0-20 30 40 50 60 70 80 90 100

15W40 Oil Temperature (C)

Figure 4-15: Friction for varying temperature at 2000 rpm for 3.OL engine

-+-Total Mechanical -A- Recip no/gas -X- Reci p gas -U- Crankshaft

200.0

180.0

160.0

140.0

120.0

100.0

80.0

60.0

40.0

20.0

0.020 30 40 50 60 70 80 90 100

15W40 Oil Temperature (C)

Figure 4-16: Friction for varying temperature at 2000 rpm for 5.4L engine

31

(U0~

5 Summary

This paper documents the changes made to the Patton Friction Model to improve the predictive accuracy

of the model. These changes were based using three engine data sets. The modifications were made

knowing that the friction work components fell under three main categories; independent of speed

(boundary lubrication), proportional to speed (hydrodynamic friction) or to speed squared (turbulent

dissipation), or some combination of these. Modifications were made to the hydrodynamic terms to

include their sensitivity to differences in oil grades and temperature. Changes to the boundary and

turbulent terms were made knowing that engine design has changed to lower the total friction losses in an

engine. Based on the model evaluation and subsequent use of the improved model, the following

conclusion were drawn:

1. The experimental data shows that engine friction has decreased as engine design has improved.

Main areas where model changes to reduce friction were piston friction and pumping losses.

2. More reliable predictions of crankshaft, reciprocating, valvetrain, and auxiliary friction and

pumping losses can be made. The crankshaft and auxiliary friction models gave noticeable

differences between model and data. The model under predicted crankshaft friction. For the

auxiliary model, a constant difference was seen at speeds higher than 3500 rpm. However, the

sum of the total friction losses gives an accurate prediction of the total engine friction.

3. The oil viscosity scaling with temperature allowed comparing the friction power loss from engine

start to hot conditions. Comparing the total mechanical friction at 20 and 90'C showed a friction

loss of about 2.1. This was comparable to the data that Shayler obtained from the firing tests.

4. The model now predicts, with acceptable accuracy, total friction and its major components in

modem gasoline engines.

32

. UM I-M-"'I II "PITim"''' T"MMP'- IIlllllllli IIII I E I IIIHoolMllllMIMmimmil NW.ie . .Mi 11 ([ l y

Bibliography

1. Wakuri, Y. et. al., Studies on Friction Characteristics of Reciprocating Engines, SAE Paper

952471.

2. Patton, K. J., Nitschke, R. G., and Heywood, J. B., Development and Evaluation of a

Performance and Efficiency Model for Spark-Ignition Engines, SAE Paper 890836

3. Heywood J.B., "Internal Combustion Engine Fundamentals" McGraw-Hill, New York, 1988,

pgs. 712-745.

4. Stribeck Diagram. http://www.egr.msu.edu/erl/case/sections/ch_07_26_02_d-3.htm

5. Stone, R.., "Introduction to Internal Combustion Engines" Society of Automotive Engineers, Inc.,

Warrandale, Pa., 1999.

6. William, J.A., "Engineering Tribology" Oxford University Press, New York, 1994, pg. 241.

7. Chevron Supreme Motor Oil. http://library.cbest.chevron.comlubes

8. Taylor, R. I., Brown, M.A., Thompson, D. M., and Bell, J. C., 1994, The Influence of Lubricant

Rheology on Friction in the Piston Ring-Pack, SAE Paper 941981.

9. Tian, T., Modeling the performance of Piston Ring-Pack in Internal Combustion Engines, PhD

Thesis, Department of Mechanical Engineering, MIT, June 1997.

10. Shayler, P.J., Christian, S.J., and Ma, T., A Model for the Investigation of Temperature, Heat

Flow and Friction Characteristics During Engine Warm-up, SAE Paper 931153.

11. Cross-sectional view of an engine. http://www.howstuffworks.com

33

Appendix

A. 1: Definition of Symbols

A = areaafmep = auxiliary friction mean effective pressure (kPa)B = bore (mm)C = radial journal bearing clearanceCr = piston roughness constantcfmep = crankshaft friction mean effective pressure (kPa)Db = bearing diameterDV = valve diameter (mm)Ap = pressure drop (kPa)

T1V = volumetric efficiencyf = friction coefficientFf = friction forcefmep = friction mean effective pressure (kPa)Fn = normal forceFt/Ft0 = piston ring tension ratioL = lengthLb = bearing length (mm)LS = piston skirt lengthL, = maximum valve lift

m = mass flow ratemfmep = mechanical friction mean effective pressure (kPa)

= oil viscosityN = engine speed (rpm)nb = number of bearingsnc = number of cylindersnv = number of valves

Pa = atmospheric pressure (kPa)pi = intake manifold pressure (kPa)Pf = friction power losspmep = pumping mean effective pressure (kPa)re = exhaust valve diameter / borer; = intake valve diameter / borerfmep = reciprocating friction mean effective pressure (kPa)

p = densityS = stroke (mm)SP = mean piston speed (m/s)tfmep = total friction mean effective pressure (kPa)V = velocity

Vd = displaced volumevfmep = valvetrain friction mean effective pressure (kPa)

34

A-1: Cross-section of an engine at intake stroke

Rm Aer M

Q voe cover

oEnoue BMack

00i Pano Oil sump

Rocker Arm& spring

o wk Mlugo Ext"St Port

oCcOnGctki Rod0Rod Beuftg

Q 'r*"""f

0-

ammC 'au ~ tww Wk~ .

~INTAKE

COMPRES8ON

COMBUSTION

o EXHAUST

Top Deed Center

A.2: Oil Grades

Oil Grade k (cSt) 91 (0 C) 02 (0 C) P. /u0 C1 c2 (*C4)

OW40 0.01341 1986.4 189.7 0.67 2.5 0.026

5W20 0.04576 1224 134.1 0.94 2.5 0.029

5W40 0.15 1018.74 125.91 0.8 2.3 0.0225

10W30 0.1403 869.72 104.4 0.76 2.3 0.0225

10W50A 0.0352 1658.88 163.54 0.49 2.43 0.0218

1OW50B 0.0507 1362.4 129.8 0.52 2.28 0.0269

15W40A 0.1223 933.46 103.89 0.9 2.3 0.0225

15W40B 0.03435 1424.3 137.2 0.79 2.5 0.026

20W50 0.0639 1255.46 117.7 0.84 2.3 0.0225

SAE10 0.0258 1345.42 144.58 1 2.3 0.0225

SAE30 0.0246 1432.29 132.94 1 2.3 0.0225

SAE50 0.0384 1349.94 115.16 1 2.3 0.0225

35

0-

A-2: Modified Expression for Total FMEP

An expression for total engine fmep, which is made up of all the component fmep terms, is:

finep = 1.22x0 B Sn+ 3.03x 10-4

ND3Lbnb

B'SnC )+ 1.35 x10' 0 DN 2 n,

+ 2.94 x10 2 KJ + 4.06X104 tEC> + +3.03 Xi1-~ 4!rND3 LbnbN )B 2 B 2 BSn~

I+4.12+6. 89 -- 0.088 rc + 0.182Pa LJ0

F rc 1.33-2KS

F )

+244 B Sn 20 B SnC.

+C 1

+Coh j 4;N;:nv.)

+ 500)In,

+ Co, (1 +500) Lvn

+6.23+5.22x10-3N - 1.79 x10~N 2

+ (Pa - Pi)+ 3.0 x 10-3&Pi J 2 P4(Pa ) nv ri

+ 0.178 fKP a

S

2

,p+ 3.0 x 10 -

K is equal to 2.38 x 10-2, and Cf, Cf, COh, Com are constants based on the valvetrain mechanism.

36

+ C4Nnv

Pa

s2

p

nvr

A.3: Sample Input Configuration for Friction Model

6 NCYL - NUMBER OF CYLINDERS2 BLKCON - 1/2: LINE/VEE BLOCK CONFIGURATION2 BLKMTL - 1/2: IRON/ALUM BLOCK MATERIAL3 VLVCON - 1/2/3: OHV/SOHC/DOHC VALVETRAIN CONFIGURATION2 FOLTYP - 1/2: FLAT/ROLLER FOLLOWER TYPE2 VLVMEK - 1/2/3: ROCKARM/FINGFOUDIRACT VALVETRAIN MECHANISM

6.0 IVO - INTAKE VALVE OPENING POINT (deg-CA BTDC)236.0 IVDUR - INTAKE VALVE OPEN DURATION (deg-CA)50.0 EVO - EXHAUST VALVE OPENING POINT (deg-CA BBDC)245.0 EVDUR - EXHAUST VALVE OPEN DURATION (deg-CA)

WOTSWITCH (WIDE OPEN THROTTLE =1)1 TUNSWITCH (TUNED=1)1 MOTSWITCH (MOTORING=1)

1 intake valves1 exhaust valves

33.8 intake valve diameter (mm)37.5 kxhaust valve diameter (mm)

*v/borev/bore

15W40 Oil Type (0W40,5W20,5W40,10W30,10W50,15W40)90 Temperature (Celsius)

14.326 Viscosity (cSt at lOOC)10.3 Reference viscosity (cSt at 90C)

Ratio of viscosity

37